The present invention relates generally to control systems for hydraulic devices and, more particularly, to such control systems for use in machine tools.
Machine tools typically require a workpiece or tool element to be positionable or movable linearly and/or rotatably during the course of machining. This positioning or movement is often controlled hydraulically. For example, hydraulically actuated pistons may be employed adjacent the sides of the main slides of lathes to provide linear motion between the head and tail ends of the bed. Four-way valves typically control the flow of hydraulic fluid to and from these pistons. Previously, electrohydraulic servo valves have been used to actuate the four-way valves.
Unfortunately, electrohydraulic servo valves are comparatively expensive and do not provide sufficiently reliable and accurate control over the valve position. Such control is desirable to achieve greater machining precision. It has often been very difficult to accurately control the null position of the four-way valve such that no fluid flow is permitted to or from hydraulic control lines, especially under changing working conditions such as increasing temperature. It has also been difficult to precisely determine the rate of fluid flow in the hydraulic control lines for a given four-way valve setting with such servo valves. At least in part, this is due to component hystereses. Sources of these hystereses include friction and inertia between moving parts. For example, friction and inertia not only retard movement of the four-way valve spool within the valve housing, but also prevent it entirely below a given level of applied motive force. Also, overlap of the outer surface of the four-way valve spool over flow ports results in a wide null region or deadband wherein the application of motive force sufficient to cause the valve to move still does not result in fluid flow-through the valve. Another disadvantage of these servo valves is it has often been necessary to provide separate control systems for linear and rotary motive systems, thus increasing machine costs and efficiency losses even more.
Stepping motors have been previously employed to control movement and positioning generally. A conventional way to drive stepping motors is by using resistors to limit the current going to the different motor windings. However, this driver method produces operating inefficiencies from the wasted power and heat build-up in the resistors used to limit stepping motor winding currents. Also, such drivers provide only full or half step resolution on particular stepping motors.
Another prior method of driving stepping motors employs a "chopper" driver which detects the current level going through the winding and limits it by shutting off the current completely when a predetermined upper limit is attained. When the current level then decays to a predetermined lower limit, the chopper driver turns the current flow on again. Thus, the current for each winding of a stepping motor is continuously turned off and on again to produce an average current over time which is within the ratings of the stepping motor. The chopper driver permits modulation or control of the amount of current going through each winding by adjustment of the levels at which the current is turned off and on. Chopper drivers are typically used in microstepping operating systems wherein the position of the stepping motor is controlled between its natural or inherent poles. However, chopper driver driver systems have been found to be impractical where extremely fine resolution is required. This results since, when the current requirement approaches zero in a particular control winding, significant induced currents arise in that winding from adjacent windings having a high current running therethrough. Thus, within practical limits, it is thus nearly impossible to achieve zero current in a particular control winding. This means that the angularity of the stepping motor cannot be precisely established and, if the stepping motor was used to actuate a valve, precise control of the valve openings and the flow rate therethrough cannot be attained. Where chopper driven stepping motors are employed with machine tools specifically, further imprecision may result because the continuous turning on and off of the current produces a continuously variable torque.
Various modern hydrostatic transmission systems also require increasingly greater precision in their control systems. For example, a variable displacement pump (VDP) may be employed to drive a fluid motor providing spindle rotation. Control for the VDP has been provided by a closed-loop analogue system where analogue command signals are input to a controller which displaces a hydraulic actuator that acts through a valve system to displace the VDP swash plate. Analogue feedback signals are provided off the fluid motor and are input to the controller. The controller makes a comparison between the command signals and the feedback signals to generate error signals to alter the swash plate displacement. These error signals gradually decrease along a continuous curve as the actual swash plate position approaches the desired position. Several problems, largely a result of VDP and fluid motor idiosyncrasies, have been found where extremely precise control of these transmission systems is desired.
There are several characteristics of individual VDPs that make control with simple closed-loop analogue systems very difficult. For example, there is often a considerable amount of deadband and hystereses in the swash plate operation itself. In other words, an increase of the command signal does not cause immediate swash plate displacement. Also, the command signal must exceed a given level before any swash plate movement results. With analogue systems, this can easily result in imprecisions and instabilities. At some point, the error signal will not be sufficient to cause swash plate displacement. Thus, optimum VDP output will never be obtained. Further, the lag time in swash plate response to command signals typically results in excessive overshooting and undershooting. This unresponsiveness and control inaccuracy is exacerabated when the transmission system must be operated at a fixed speed under varying load conditions. Increasing the gain of the near zero error signal to compensate for the hystereses often does not achieve greater accuracy. Instead, because of the swash plate response lag time, the swash plate displacement oscillates about the desired position for an extended period of time.
Another problem encountered in prior hydrostatic transmission systems is when abrupt and large speed changes are necessary. It is often desirable to go from a low speed to a high speed or from one direction to another instantaneously or at least to accelerate in the shortest possible time. Where analogue loop command systems have been employed, the initial swash plate displacement is relatively fast because the error between the command and feedback signal is quite large. However, as the optimum level is approached along a continuous error signal curve, the error becomes less and less and will eventually cease to be sufficient to drive the swash plate such that the optimum is not attained. Inputting a larger command signal may cause the optimum swash plate displacement to be exceeded, but then the same hystereses problem results when the swash plate position is backed down to the optimum level.
It is therefore an object of the present invention to provide an improved motive control system having increased precision and responsiveness.
Another object is the provision of an improved hydraulic control system for accurately actuating valve elements and control flow therethrough.
A further object is to provide a control system for hydraulic actuating systems which compensates for various system component characteristics as well as operating conditions.
Yet another object is the provision of a control system for machine tools having increased response speed toward achieving desired changes in movable element orientation.
These and other objects of the present invention are provided in a control system for a hydraulic actuating system having a multi-phase stepping motor with a driver therefor which supplies and controls DC current through each of the motor windings such that no current is induced into the control winding from adjacent windings. The stepping motor precisely controls the rotation of an eccentric cam engaging an actuating plunger which actuates a control valve of the hydraulic actuating system. The driver receives control signals from a computer programmed to calculate, from user input command signals indicative of the desired system condition, and generate those control signals which most quickly achieve a rough approximation of the desired system condition. After this rough approximation is attained, the computer generates new control signals for the driver as a function of feedback signals and/or system component characteristics so as to precisely achieve the desired system condition.
The control system of the present invention has been found to be particularly suitable for use in numerical control of machine tools, such as lathes, having either for controlling workpiece or cutting tool motion. Control programming of the computer permits extensive customization without loss of precision.
Further objects, features, and advantages of the present invention will become more apparent from the following description when taken with the accompanying drawings which show, for purposes of illustration only, several embodiments in accordance with the present invention.